Internally cooled internal combustion engine and method thereof

ABSTRACT

An internal combustion engine is equipped with a water injector for cooling the internal combustion engine by a spray of atomized water into the intake track or combustion chamber prior to ignition. The atomized water spray may be in the intake manifold or directly in the cylinder. The water is injected at a volume of between a ratio of about 95% fuel to about 5% water and about 50% fuel and about 50% water. The temperature of the internal combustion engine is maintained at between about 95° C. and about 200° C. during operation.

CROSS-REFERENCE TO RELATED APPLICATIONS

The instant application claims priority from U.S. Provisional PatentApplication No. 61/753,719 filed Jan. 17, 2013, the contents of which isincorporated herein by reference. Additionally, the instant applicationclaims priority from PCT Application No. PCT/IB2013/002593 filed on Nov.20, 2013, the contents of which are incorporated herein by reference.

FIELD OF THE INVENTION

The present invention relates generally to internal combustion engines.More specifically, the present invention relates to an internalcombustion engines with exhaust gases recirculation.

This invention pertains to internal combustion engines with at least onereciprocating piston that operate with directly cooled exhaust gasesrecirculation (EGR). The principles set forth herein can be used in bothspark-ignition (SI) engines typically operating on gasoline (petrol),natural gas, or ethanol blends, or in compression-ignition (CI) enginestypically operating on diesel, biodiesel, JP-8 or other jet fuelvariants, kerosene, or heavy oil. This invention is applicable to bothnaturally aspirated and forced aspiration internal combustion engineswith exhaust gases recirculation. This invention is applicable in directfuel injection and port fuel injected engines.

BACKGROUND

The use of EGR in internal combustion engines is a well understood andwidely applied in commercial products. Exhaust gases recirculated to thecombustion chamber of a gasoline engine displace the amount ofcombustible charge in the cylinder, and in a diesel engine the exhaustgases displace excess oxygen in the pre-combustion mixture. Thedisplacement of combustible charge results in a lower combustiontemperature and is effective in reducing the formation of NOx whichforms primarily when a mixture of nitrogen and oxygen is subject totemperatures above 1371° C. (1644° K.). Recirculated exhaust gasesdisplace intake air and decrease the charge density through heating.These combined effects contribute to reducing pumping losses resultingin an increase in engine efficiency, albeit at reduced power. EGR istherefore an effective method for reducing Nitrogen oxides (“NOx”)emissions in both SI and CI engines, as well as improving Otto-cycleengine efficiency.

The reintroduction of exhaust gases back into the combustion chamberreduces peak combustion temperatures. This reduction in temperature islargely because the returned exhaust gases do not participate in thecombustion and thus delivers no combustion energy. The exhaust gasesprovide additional thermal mass and allow combustion energy todistribute to a higher overall thermal mass, where the product of massand heat capacity (m*Cv) is higher with EGR than without EGR. Thetemperature reduction provided by EGR recirculation reduces combustiontemperature and is therefore effective in controlling and reducing NOxformation. EGR allows for higher manifold pressures at any given load,resulting in a reduction in charge cycle work, lowering fuelconsumption.

There are two methods of re-routing exhaust gases back into thecombustion chamber. The first method is internal exhaust gasesrecirculation (i-EGR) via valve phasing or valve overlap. Valve overlapis the condition in which the intake valve is opened early to allowexhaust gases to enter the intake track during the exhaust stroke or thecondition in which the exhaust valve is kept open late during the intakestroke to allow exhaust gases to return to the combustion chamber. Thisis commonly achieved by utilizing variable valve timing systems to varythe camshaft phasing to adjust the valve event according to the engineoperating point to optimize the EGR benefits. This is illustrated inFIG. 1, showing a schematic of a prior art engine showing the flow ofexhaust gases with internal EGR via intake valve 11, which is openedearly at the end of the exhaust stroke to allow exhaust gases 14 fromthe combustion chamber 15 to enter the intake track 16 during the piston12 exhaust stroke and mix with intake charge air 13 entering thecombustion chamber 15 during the piston 12 intake stroke.

Referring to FIG. 2, a schematic of a prior art engine shows the flow ofexhaust gases with internal EGR via exhaust valve 21. The exhaust valve21 remains open late after the piston 22 exhaust stroke, and during theintake stroke of the piston 22 to allow exhaust gases 23 in the exhausttrack 26 to return 24 to the combustion chamber 25.

The second method of exhaust gas recirculation is via an exhaust gasloop external to the combustion chamber which may or may not comprisecorresponding controlled EGR valves (e-EGR). The EGR valve is activatedelectronically depending on the engine operating point to feed theappropriate amount of exhaust gases back into the fresh intake air-fuelmixture. FIG. 3 shows a prior art schematic of an engine in which EGR isprovided via an external loop with an external EGR cooler 34. Theexhaust gas 31 is expelled from the combustion chamber 33 during thepiston 32 exhaust stroke. The exhaust gas 31 is channeled from theexhaust track 37, by means of tubes, pipes, channels, and other means toan external heat transfer device 34, in the form of a heat exchanger orlike embodiment to cool the exhaust gases. The cooled exhaust gases 35are channeled from the heat transfer device 34 into the intake air flow36 prior to or within the intake air track 38. As previously described,the additional EGR gas increases the thermal mass of the mixed intakecharge.

Both solutions have drawbacks. With e-EGR, a time delay is introducedbetween an EGR percentage request by the engine management system andthe exhaust gases arrival at the engine inlet. This delay causes controlissues which lead to lower engine efficiency. With i-EGR control isimproved, but very high gas temperatures are re-circulated, leading to aloss of volumetric efficiency and a limitation on how much EGR can beachieved prior to knock onset. Industry and academic work has beenperformed on cooled EGR utilizing an external heat exchanger to cool theexhaust gases and all focus has been on external EGR loop coolingbecause it has been the most effective and feasible method to implementcooled EGR.

The emissions reduction potential of EGR systems can be improved furtherthrough cooled EGR systems. Cooled EGR is widely utilized in compressionignition engines, where the EGR system is integrated into the highpressure exhaust and charge loop of a turbocharged diesel engine. Theexhaust gases are recirculated from the main exhaust flow between thecylinder and the exhaust gases turbine. The exhaust gases pass throughan intercooler or heat exchanger, which utilizes a secondary externalcooling source, to transfer heat from the exhaust gases though a solidmedium in the form of a heat exchanger. The cooled exhaust gases arethen introduced into the intake air loop of the engine, either in thehigh pressure loop between the compressor and the cylinder or in the lowpressure loop upstream of the compressor.

A cooled external EGR system may use a valve to regulate the volume ofre-circulated exhaust gases controlled by the engine management system,the exhaust pipes, the exhaust gas cooler and the intake pipes. Thesesystems utilize an external cooling agent through a form of heatexchanger in order to extract heat from the hot exhaust gases prior tointroducing the exhaust gases into the cylinder chamber. Cooled EGRsystems expose the exhaust gas cooler to an extreme temperature up toabout 450° C. in passenger cars and about 700° C. in commercial vehicleapplications.

SUMMARY OF THE INVENTION

An embodiment of the present invention is an internal combustion engineincluding: at least one cylinder. Each cylinder has a combustionchamber, a piston, an air intake valve, and an exhaust valve. An airintake track is provided in communication with each air intake valve,and an exhaust track is provided in communication with each exhaustvalve. A fuel handling system, with at least one fuel injector, isconfigured to inject fuel into the combustion chamber or intake track.An ignition system is configured to ignite the fuel in the combustionchamber at an end portion of a compression stroke of the piston.Additionally, the present invention includes a primary cooling systemfor maintaining the internal combustion engine within a predeterminedoperating temperature range. The primary cooling system includes a waterreservoir in fluid communication with an injector provided to cool theinternal combustion engine. The injector is arranged to inject acontrolled amount of liquid water into the combustion chamber or intaketrack.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic of a section of a prior art engine showing theflow of exhaust gases with internal EGR via intake valve.

FIG. 2 is a schematic of a section of a prior art engine showing theflow of exhaust gases with internal EGR via exhaust valve.

FIG. 3 is a schematic of a section of a prior art engine showing theflow of exhaust gases with external EGR via an external loop with EGRcooler.

FIG. 4 is a schematic of a naturally aspirated internal combustionengine of the present invention with direct fuel injection and enginesystems showing internal EGR, via intake or exhaust valves, with directEGR cooling via a water injector directly into the combustion chamber.

FIG. 5 is a schematic of a naturally aspirated internal combustionengine of the present invention with direct fuel injection and enginesystems showing the flow of exhaust gases through an external EGR loopwith direct EGR cooling via a water injector directly into thecombustion chamber.

FIG. 6 is a schematic of a naturally aspirated internal combustionengine of the present invention with port fuel injection and enginesystems showing the flow of exhaust gases through an external EGR loopwith direct EGR cooling via water injector directly into the combustionchamber.

FIG. 7 is a schematic of a naturally aspirated internal combustionengine of the present invention with port fuel injection and enginesystems showing the flow of exhaust gases through an external EGR loopwith direct EGR cooling via water injection in the intake track.

FIG. 8 is a schematic of a turbo charged internal combustion engine ofthe present invention with direct fuel injection and engine systemsshowing the flow of exhaust gases through an external EGR loop, highpressure and low pressure, with direct EGR cooling via a water injectordirectly into the combustion chamber.

FIG. 9 is a schematic of a turbo charged internal combustion engine ofthe present invention with port fuel injection and engine systemsshowing the flow of exhaust gases through an external EGR loop, highpressure and low pressure, with direct EGR cooling via a water injectordirectly into the combustion chamber.

FIG. 10 is a flow diagram of a control process performed by anembodiment of the present invention.

FIG. 11 is a schematic of a naturally aspirated internal combustionengine of the present invention with direct fuel injection with internalcooling via a water injector directly into the combustion chamber.

FIG. 12 is a block representation of an internal combustion engine ofthe present invention with internal cooling and water recovery system.

DETAILED DESCRIPTION

The present invention provides a four-stroke spark ignition orcompression ignition (diesel) internal combustion engine that operatesat substantially higher thermodynamic efficiency than conventionalengines through the use of lean fuel mixtures, high compression ratios,higher operating temperatures, exhaust gases recirculation (EGR), andwater injection in the EGR path, intake manifold or cylinder.

In the context of the present invention, the term “intake track” refersto any part of the fresh air path between the environment, i.e., the airintake, and the combustion chamber. Thus, the intake track includes theair intake, air inlet, any fresh air conduit, and the intake manifold.In the context of the present invention, the term “exhaust track” refersto any part of the exhaust gases pathway including, for example, thecylinder outlet, the exhaust manifold, any exhaust gases conduit andconnections, and may include a muffler and exhaust pipe, venting fumesto the environment. The term “EGR track” refers to any part of anexhaust gases recirculation system between a shunt in the exhaust trackthat diverts a portion of the exhaust gases to the EGR system, and anyconduit, valves, connections, or other parts of the EGR system thatdefine a path for recirculated exhaust gases, until the EGR gases areintroduced into the intake track.

As used herein the term “λ” refers to the stoichiometric ratio of oxygenin air to fuel. Stoichiometric ratio of oxygen in air to fuel meansthere is one mole of oxygen (in air) for each mole of carbon in ahydrocarbon fuel and one mole of oxygen for every two moles of hydrogenin fuel. This stoichiometry translates to a weight ratio of about 14.7:1(w/w, air:gasoline) for gasoline. Higher λ values indicate leanermixtures, or more air per unit of fuel. Thus, λ greater than 1 means aratio (for gasoline) of greater than 14.7:1 w/w. Different fuel typesrequire different stoichiometries. For example, stoichiometric air tofuel for methanol is about 6.5:1, ethanol is about 9.0:1, diesel is14.4:1, natural gas is 16.6:1, and methane is 17.2:1.

Conventional internal combustion engines equipped with exhaust gasrecirculation provide a heat exchanger, such as a radiator in theexhaust gases' recirculation path in order to cool the exhaust gasesprior to reintroduction of the exhaust gases into the combustionchamber. In contrast, the inventive engines disclosed herein do notrequire a heat exchanger at all, thereby minimizing heat losses to theenvironment. Internal temperature control and engine cooling in thepresent invention is provided by the lean fuel mixtures, EGR, and waterinjection either into the intake manifold or directly into the cylindersof the engine. Accordingly, the inventive engines have been shown tooperate at up to 50% thermodynamic efficiency. Nevertheless, in anembodiment a heat exchanger may be utilized.

Conventional Otto-cycle engines are limited to compression ratios of nomore than 12:1 when using high octane fuels in a spark ignition engine,and no more than 23:1 in compression ignition engines. Compressionratios greater those noted above are generally understood to causeengine damage by, for example, inducing premature detonation of the fuelin the combustion chamber, and to suffer from excessive heat losses.However, high compression, when the cylinder pressure can be properlycontrolled has the benefit of increased efficiency in converting thecombustion of fuel to mechanical energy.

Conventionally, EGR cooling is recognized as desirable to minimizepumping losses, control engine temperature, and minimize NOx production.In an embodiment of this invention, EGR gases are cooled internallywithout the need for an external heat exchanger (EGR cooler). Theinventive methods allow much higher amounts of EGR to be utilizedwithout a knock-limit penalty, without reduced charge density andwithout volumetric efficiency losses. This may be most effective oninternal EGR loops, though the invention can be used with any method ofEGR recirculation, and for both turbocharged and non-turbochargedengines, as well as both port fuel injected and direct fuel injectedengines.

Conventionally, external EGR cooling is commonly employed. The inventiveengines are designed to run at higher internal temperatures thanconventional engines, which are made possible by the lean fuel mixtures,EGR systems, and internal cooling of the EGR with water. The phasetransition of the water from liquid to vapor consumes heat energypresent in the recirculated exhaust gases, thereby lowering thetemperature of the recirculated exhaust gases to a temperature that islower than the temperature of the re-circulated exhaust gases prior tointroduction.

Cooling of the recirculated exhaust gases in the present invention,thus, occurs at one or more positions along the EGR track and intaketrack by way of a spray of atomized water directly into the recirculatedexhaust gases. Thus, in the case where the exhaust gas is cooled afterbeing introduced into the intake track, for example, the recirculatedexhaust gas has essentially the same temperature at the point just priorto cooling in the intake track as at the exhaust manifold.

In an embodiment, the inventive EGR includes a water reservoir, a waterhandling system comprised of pipes or tubes and a rigid distributionrail, and one or more water injector(s), and a computer control systemthat uses a reference table to inject varying amounts of water inresponse to the engine load, speed and current EGR conditions.

The water can be injected into the engine either at the air intake port(port injection) or directly into the combustion chamber (directinjection). Direct injection is the preferred embodiment as it allowsmore accurate and precise control over the water spray timing andposition when compared to port injection.

This system can be used with any internal combustion engine employingEGR; either two or four stroke, and fueled by a combustible liquid usedfor fuel, such as gasoline, diesel, ethanol, methanol, hydrogen, naturalgas, or a mixture thereof, and with spark or compression ignitionengines. The example embodiments discussed herein are of four strokeengines using either spark or compression ignition. However, based onthe disclosure provided herein, one of ordinary skill in the art willreadily appreciate the alterations and modifications necessary to applythe present invention to two stroke engines as well as other forms ofreciprocating internal combustion engines.

In an embodiment, an internal combustion engine is provided, operatingon a fuel, such as hydrocarbon fuel, with internally cooled exhaustgases recirculation, with at least one cylinder and a reciprocatingpiston therein, a combustion chamber in the cylinder, an air intakemanifold with at least one air intake valve, an exhaust manifold with atleast one exhaust valve, a fuel handling system with a fuel injector,and an ignition system; wherein the engine has a mechanical compressionratio greater than 12:1 and less than 40:1, and operates at an air tofuel ratio of λ greater than 1 and less than 7.0; wherein the engine hasmeans to recirculate exhaust gases internally or externally; wherein theengine internally cools the recirculated exhaust gases by direct contactwith predetermined quantity of atomized water injected into the exhaustgases without the use of a mixed medium heat exchanger that chills therecirculated exhaust gases.

Lean fuel mixtures are desirable in order to reduce throttling lossresulting from having to operate the engine with a partially closedthrottle as occurs when the engine is operating at a steady speed.However, leaner fuel mixtures can burn hotter in a specific range of λgreater than 1, which can result in increased emissions of NO_(x) at λgreater than 1. Operating an internal combustion engine with a leanmixture can quickly result in combustion chamber temperatures exceeding2500° F. In addition to increasing NO_(x) production, the excessivelyhigh temperature in the combustion chamber can lead to prematuredetonation of the fuel (knocking) and warping of the various componentsof the engine.

In an embodiment, a method of operating an internal combustion engine isprovided, wherein the engine uses a fuel, such as hydrocarbon fuel, withinternally cooled exhaust gases recirculation, with at least onecylinder and a reciprocating piston therein, a combustion chamber in thecylinder, an air intake manifold with at least one air intake valve, anexhaust manifold with at least one exhaust valve, a fuel handling systemwith a fuel injector, and an ignition system. The engine has amechanical compression ratio greater than 12:1 and less than 40:1, andoperates at an air to fuel ratio of λ greater than 1 and less than 7.0.Additionally, the engine has means to recirculate exhaust gasesinternally or externally, and internally cools the recirculated exhaustgases by direct contact with predetermined quantity of atomized waterinjected into the exhaust gases without the use of a mixed medium heatexchanger that chills the recirculated exhaust gases. In anotherembodiment, a method of cooling EGR gases in an internal combustionengine is provided.

The optimum λ for the inventive engines depends on the ignition type.For spark-ignition engines running gasoline, gasoline blends (forexample, with ethanol), or natural gas (primarily methane), λ will be inthe range of greater than 1 to a maximum of about 3.0. In alternativeembodiments, λ in spark ignition engines according to this inventionwill be in a range of from about 1.2 to about 2.8, or about 1.2 to about2.3, or about 1.5 to about 2.0, or about 1.5, or about 1.75, or about2.0. For compression ignition engines (diesels), λ will be in the rangeof greater than 1 to a maximum of about 7.0. In alternative embodiments,λ in the inventive engines will be in a range of from about 1.4 to about6.0, or about 1.5 to about 5.0, or about 2.0 to 4.0, or about 1.5, orabout 2.0, or about 2.5, or about 3.0, or about 3.5, or about 4.0.

The optimum compression ratio for the inventive engines depends on theignition type. For spark-ignition engines running gasoline, gasolineblends, or natural gas, conventional engines have a typical compressionratio of 10:1, with a maximum compression ratio of about 12:1 usinghigher octane fuels. These compression ratio limits are required inorder to control engine knock that would otherwise occur at highercompression ratios. By using higher compression ratios than conventionalengines, the inventive engines have the benefit of superiorthermodynamic efficiency according to the Otto cycle, in whichthermodynamic efficiency is a function of compression ratio.

The compression ratio of the inventive engines in spark-ignition mode isin the range of greater than 12:1 to about 20:1. In alternativeembodiments, the compression ratio is 13:1 to about 18:1, or about 14:1to 16:1, or about 14:1, or about 15:1, or about 16:1, or about 18:1. Forcompression ignition engines, the compression ratio will be from about14:1 to about 40:1. In alternative embodiments, the compression ratio isin a range of about 14:1 to about 30:1, or about 15:1 to about 25:1, orabout 16:1 to about 20:1, or about 16:1, or about 17:1, or about 18:1,or about 19:1, or about 20:1, or about 21:1, or about 22:1.

As noted above, internal combustion engines using spark ignition aregenerally limited to compression ratios of no more than 12:1 in order toavoid premature detonation. Thus, usage of compression ratios above12:1, as in the present invention, is not obvious given the generalknowledge of internal combustion engines. The present invention avoidsthe dangers associated with compression ratios higher than 12:1 by theuse of internally cooled EGR.

EGR is well known to provide several benefits to internal combustionengines and is commonly used. However, a shortcoming of EGR is theaddition of excess heat into the combustion chamber, which tends toincreased premature ignition (knock) and may increase NO_(x) emissions,which are dependent on combustion temperature. Consequently, atomizedwater is sprayed directly into the EGR track or intake track in theinventive engine to cool the reintroduced exhaust gases to a controlledtemperature.

Because the water-cooled EGR reduces the temperature within thecombustion chamber, a significantly leaner fuel mixture can be usedwithout producing elevated NO_(x) emissions or knocking. The leaner fuelis the second feature that makes the high compression ratios possible inthe present invention.

The amount of water injected is a function of the fuel flow and theamount of EGR employed. Fuel flow in modern engines is typicallydetermined from a mass air flow sensor or a manifold pressure sensor,which provides data to an engine control computer that determines thequantity of fuel fed to the fuel injectors. The quantity of EGR gasesshunted back into the engine is also controlled by the engine controlcomputer. In the case of external EGR, the amount of EGR is controlledby the EGR valve. In internal EGR embodiments, the valve timing isindependently controllable with variable valve timing, for example withcam phasing. Other multipliers are typically used by an engine controlcomputer to control fuel flow and EGR include engine load, intake airtemperature, exhaust oxygen sensor, and engine rpm. In the inventiveengines, the water flow will be determined by the computer using thesame parameters.

The amount of water injected can be expressed as a percentage by weightof the EGR gases injected into a cylinder prior to ignition. In anembodiment, the amount of water injected is about 10% to about 125% ofthe recirculated exhaust gases (EGR) by weight (w/w). In an embodiment,the amount of water injected is about 10% to about 100% of the EGR w/w,or about 25% to about 100% of the EGR w/w, or about 20% to about 100% ofthe EGR w/w, or about 75% to about 125% of the EGR w/w, or about 25%w/w, or about 50% w/w, or about 75% w/w, or about 100% w/w.

The amount of water injected in the inventive engines may be reducedcompared to prior art water injector embodiments, without reducing theamount of water or water vapor in the cylinder during ignition, becauseEGR gases contain substantial amounts of water vapor, since water is acombustion product of hydrocarbon fuels. Because the EGR gases are nottreated or cooled in the inventive engines (in contrast to conventionalEGR systems), the full load of water vapor in the EGR gases will becirculated back to the engine. In one aspect, this feature of theinventive EGR systems will reduce the amount of liquid water necessaryfor injection into the engine that must be carried on board a vehicle(for an engine in a vehicle) at any given moment.

The water may be injected with an injector adapted to injecting liquidsinto an engine intake manifold or cylinder. In an embodiment, a waterinjector may inject an atomized spray of water into the intake manifoldin the presence of EGR gases prior to being drawn or injected into thecylinder prior to ignition. In an embodiment, a water injector mayinject an atomized spray of water directly into the cylinder, after EGRgases have been injected or drawn into the cylinder.

The phrase “internally cooled exhaust gases recirculation”, asunderstood in the context of the present invention, is intended to meanthat no mixed medium heat exchanger is employed in the EGR track. Thus,in an engine employing internally cooled exhaust gases recirculation,there is no heat exchanger, radiator, cooling coils, jacketed cooling,air cooling fins, or other external cooling apparatus in the EGR track.The EGR track, within the context of the present invention, is definedas the exhaust gases path between the points where a portion of theexhaust gases are diverted from the exhaust track to the injection ofthe diverted exhaust gases into the intake track.

By contrast, EGR cooling with a heat exchanger is well known in theprior art. In accordance with the present invention, the only cooling ofEGR gases is from internal cooling by water directly injected into theEGR track, the intake track after injection of EGR gases, or by directinjection of water into the cylinder after EGR gases are introducedtherein.

The predetermined quantity of atomized water injected into the exhaustgases need not be pure water. In an embodiment, the water may include alower alkanol, especially a C₁ to C₄ alcohol, for example, methanol,ethanol, n-propanol, isopropanol, or any isomer of butyl alcohol. Theuse of a solution of an alcohol in water may be, for example, to depressthe melting point of the water for EGR cooling in cold climates. Forexample, a 30% mixture of ethanol in water (w/w) has a melting/freezingpoint of −20° C.

An internal EGR embodiment of this invention is illustrated in FIG. 4,showing a schematic of a naturally aspirated internal combustion enginewith direct fuel injection and engine systems showing internal EGR. Withinternal EGR, no external exhaust gas recirculation path is provided.Rather, exhaust gases are recirculated “internally” with valve phasingor valve overlap, with direct EGR cooling via a water injector directlyinto the combustion chamber. In this embodiment, the timing of intakevalve 5 or exhaust valve 6 must be independently computer controlled toprovide valve phasing or valve overlap EGR.

The operation of the internal combustion engine of the presentinvention, shown in FIG. 4, conforms generally to a standard four strokeengine. An air intake valve 5 opens at the beginning of an intake strokeof the piston to allow air to flow into the combustion chamber 1. Theair intake valve 5 closes prior to the initiation of the compressionstroke in which the piston 2 compresses the air and fuel in thecombustion chamber 1. Shortly before the top of the piston 2 travel,i.e., top-dead-center (TDC), the ignition system, i.e., spark plug 24ignites the fuel/air mixture in the combustion chamber 1. After thepiston cycle past TDC, the ignited fuel pushes the cylinder downward inthe power stroke to turn a crankshaft 26. When the piston has reachedits lowest point of travel in the cylinder during the power stroke,i.e., bottom-dead-center (BDC), the internal combustion engine beginsthe exhaust stroke. In the exhaust stroke the exhaust valve 4 opens andthe upward travel of the piston 2 forces the exhaust gases out of thecombustion chamber 1

In this internal EGR embodiment, exhaust gases are internallyrecirculated through valve phasing or valve overlap, by specialsequencing of exhaust valve 4 or intake valve 5. For example, the intakevalve may open during part of the exhaust stroke to admit some exhaustgases into the intake manifold. These gases are then recirculated backinto the cylinder during the intake stroke. In another embodiment, theexhaust valve may be opened during the intake stroke, thereby admittingsome of the exhaust gases in the exhaust manifold to the cylinder. Thus,in the embodiment of FIG. 4, one or both of the intake and exhaustvalves must be independently controlled to effect the necessary valvephasing.

As shown in FIG. 4, water from reservoir 8 is pressurized by pump 9 andis injected through injector 7 directly into the combustion chamber 1 tocool the rebreathed exhaust gases. The amount of water injected isdetermined and controlled by engine control computer 30. Also depictedin FIG. 4 is a fuel reservoir 21, a fuel pump 20, a fuel injector 12, acoil 23 and a piston rod 25.

The engine control computer 30 has connections to manifold pressuresensor 29, water pump 9, fuel pump 20, and variable valve timingcontrols (not shown). An embodiment of the engine depicted in FIG. 4operates at the high compression ratios, lean fuel mixtures, andpredetermined amount of injected water in accordance with thisinvention.

Another embodiment of this invention is described in FIG. 5, showing aschematic of a naturally aspirated internal combustion engine withdirect fuel injection and engine systems showing the flow of exhaustgases through an external EGR loop with direct EGR cooling via waterinjector 7, which injects water directly into the combustion chamber 1.Thus, exhaust gases from high compression combustion chamber 1 exitduring the exhaust stroke of high compression piston 2 into the exhausttrack 3. EGR valve 10, controlled by engine control computer 30, allowsa controlled amount of exhaust gases to enter the EGR track 11 to bedelivered to the intake track 6 without passing through an external heatexchanger. The recirculated exhaust gas temperature is higher than theintake air charge temperature.

Water injector 7 injects a predetermined amount of water into thecombustion chamber with the recirculated exhaust gases from EGR track 11and with fuel injected directly into the combustion chamber throughinjector 12. The water injected into the chamber reduces the elevatedtemperature of the recirculated exhaust gases directly in accordancewith this invention. Also depicted in FIG. 5 is engine control computer30 with connections to manifold pressure sensor 29, water pump 9, fuelpump 20, and EGR valve 10. An embodiment of the engine depicted in FIG.5 operates at the high compression ratios, and lean fuel mixtures inaccordance with this invention.

Another embodiment is shown in FIG. 6, illustrating a schematic of anaturally aspirated internal combustion engine with port fuel injection,direct water injection, and engine systems showing the flow of exhaustgases through an external EGR loop with direct EGR cooling via waterinjector directly into the combustion chamber. Exhaust gases from highcompression combustion chamber 1 exit during the exhaust stroke of thehigh compression piston 2 into the exhaust track 3. EGR valve 10 allowsan amount of exhaust gases to enter the EGR track 11 to be delivered tothe intake track 6 without passing through an external heat exchanger.The recirculated exhaust gas temperature is higher than intake aircharge prior to water injection. In this embodiment, fuel is injectedinto the intake track (port injection), rather than directly into thecylinder, through fuel injector 12.

Water injector 7 injects a specific and controlled amount of waterdirectly into the combustion chamber with the recirculated exhaust gasesfrom EGR track 11 and cools the elevated gas temperature prior toignition.

Another embodiment is shown in FIG. 7, illustrating a schematic of anaturally aspirated internal combustion engine with port fuel injectionand port water injection. Engine systems are shown directing a flow ofexhaust gases through an external EGR loop with EGR cooling via waterinjection in the intake track. Exhaust gases from high compressionchamber 1 exit during the exhaust stroke of the high compression piston2 into the exhaust track 3. EGR valve 10 allows a controlled amount ofexhaust gases to enter the EGR track 11 to be delivered to the intaketrack 6 without passing through an external heat exchanger. Therecirculated exhaust gases admitted to intake track 6 have a greatertemperature than the intake air. The EGR gases are cooled by water fromreservoir 8 pressurized through pump 9 and injected into the intaketrack by injector 7. Gases with fresh air, cooled EGR gases, watervapor, and fuel are aspirated into combustion chamber 1 during theintake stroke. The engine control computer, sensors, and relatedconnections are omitted for brevity from FIG. 7.

Another embodiment is shown in FIG. 8 illustrating a schematic of aturbocharged internal combustion engine with direct fuel injection,direct water injection, and external EGR. Engine systems are showndirecting the flow of exhaust gases through an external EGR loop, whichmay be either a high pressure loop 11 or a low pressure loop 17, orboth. In this embodiment, exhaust gases following ignition from highcompression chamber 1 exit to exhaust track 3 during the exhaust strokeof the high compression piston 2. The engine exhaust in this embodimentdrives turbine 14, which is connected to compressor 13 that pressurizesfresh air 15 from air intake path 28 and other gases in intake manifold6. In a high pressure EGR bypass 11, exhaust gases from exhaust pipe 3are shunted to the intake manifold before turbine 14. EGR valve 10,under computer control as described above, controls the amount ofexhaust gases entering the EGR bypass 11 to be delivered to the highpressure intake track 6.

Accordingly, the EGR gases enter the intake manifold 6 without passingthrough an external heat exchanger, which provide recirculated exhaustgases temperature at higher than the intake air charge temperature. Inthe case of the low pressure EGR loop, a portion of exhaust stream 16,after exiting the turbocharger turbine 14, is shunted to air intake intothe fresh air intake 28 through EGR bypass 17 controlled by valve 18.

Water from reservoir 8 is pressurized by pump 9 and fed to injector 7 toinject a controlled amount of water directly into the combustion chamber(1) containing the recirculated exhaust gases and with fuel injecteddirectly into the combustion chamber through injector 12. The waterinjected into the chamber 1 reduces the elevated temperature of therecirculated exhaust gases directly. The engine control computer,sensors, and related connections are omitted for brevity from FIG. 8.

Another embodiment is shown in FIG. 9, illustrating a schematic of aturbocharged internal combustion engine with port fuel injection, directwater injection, and engine systems showing the flow of exhaust gasesthrough an external EGR loop, high pressure and low pressure, withdirect EGR cooling via a water injector in the intake manifold 6. Thisembodiment is similar in operation to the turbocharged embodiment ofFIG. 8, with high and low EGR bypass embodiments, but with port fuelinjection rather than direct fuel injection.

In another embodiment (not shown), a turbocharged engine can employ theinventive EGR and water injection, with port fuel and port waterinjection. In another embodiment, a supercharger is used. By the term“turbocharger” is meant an air compressor driven by exhaust gases. Bythe term “supercharger” is meant an air compressor driven by amechanical linkage to the engine.

In other embodiments, the embodiments illustrated in FIGS. 4-9 can beused with compression ignition engines, but without the spark ignitionsystem.

Table 1 shows experimental results of a VW 1.9 L 4 cylinder turbochargeddirect injection diesel engine, with 19:1 compression ratio, andexternal EGR modified to include a water injector in each cylinder. Thevaries depending on engine load, but was never less than 1.1, and rangedup to about 1.5 in this test engine. EGR and were inverselyproportional, so that at higher λ, EGR was reduced. EGR was varied from0% to 30%. Water was varied from 0% to 100%. The highest operatingefficiencies (rows 17-21) had elevated NOx production. Increasing thewater amount or EGR amount decreased NOx production significantly withminimal effect on overall efficiency, as shown in experiments 5, 11, 21and 23.

TABLE 1 Experimental results with a four cylinder diesel engine. SpeedEGR BSFC BSFC λ NOx No. BMEP RPM Rate % Water/Fuel % g/kWh g/kWh ηFc λ %ηFc % (ppm) 1 6BAR 1800 0 25 233.1 236.7 35.2 35.7 451 2 6BAR 1800 0 50233.1 240.5 34.6 35.7 380 3 6BAR 1800 0 100 234.8 246.7 33.7 35.5 317 46BAR 1800 10 25 233.1 236.8 35.1 35.7 462 5 6BAR 1800 10 50 233.4 240.534.6 35.7 413 6 6BAR 1800 10 100 232.8 242.6 34.3 35.8 321.8 7 6BAR 180020 25 234.7 231.4 36 35.5 242.9 8 6BAR 1800 20 50 234.7 232 35.9 35.5191 9 6BAR 1800 20 100 237.9 237.5 35.1 35 142.8 10 6BAR 1800 30 25247.1 241.9 34.3 33.7 70.5 11 6BAR 1800 30 50 248.7 242.4 34.4 33.5 57.312 6BAR 1800 30 100 252.9 245.6 33.9 32.9 38.4 13 6BAR 1800 0 0 231233.7 35.6 36 474 14 6BAR 1800 35 25 272.1 258.5 32.3 30.7 51.75 15 6BAR1800 0 0 235.1 235.2 35.4 35.4 472.9 16 12BAR 2000 0 0 209.4 205.3 39.740.5 1663 17 12BAR 2000 0 25 209.6 209 39.7 39.8 1492 18 12BAR 2000 0 50210.2 209.4 39.6 39.7 1291 19 12BAR 2000 0 100 211.3 215 39.4 38.7 123120 12BAR 2000 10 25 210.8 209 39.5 39.8 1195 21 12BAR 2000 10 50 210.6208.9 39.5 39.8 1094 22 12BAR 2000 10 100 211.1 215 39.4 38.7 621 2312BAR 2000 20 25 215.5 214 38.6 38.8 374 24 12BAR 2000 20 50 215.8 216.638.6 38.4 403

The engine test results in Table 1 show a maximum efficiency of 39.5%with 10% EGR and 25% or 50% water injection (experiments 20 and 21).

In the present invention, the amount of atomized water:air:fuel mixture,and the amount of EGR employed at any given time is controlled by anengine controller (ECU). Specifically, the engine controller receivessignals relating to position of the accelerator, exhaust temperature,vehicle velocity, valve timing and position, air:fuel ratio, forexample. These signals are generated by respective sensors, as wellknown in the art and provided electronically to the engine controller.The signals provide the control parameters for adjusting the amount ofEGR, as well as the amount of atomized water injected into the EGR trackto attain a desired temperature of the recirculated exhaust gases. Inaddition, the air:fuel mixture is adjusted based on the above signals tooptimize the power output and minimize throttling loss during engineidle and cruising conditions.

In situations where a vehicle employing the inventive engine iscruising, the air:fuel mixture is at its leanest. However, this createsa significant amount of heat within the combustion chamber, as explainedpreviously. Thus, the EGR is cooled to a lower temperature byintroducing a greater volume of atomized water into the EGR track. Inthis way the compression ratio can be kept high and the air:fuel ratiocan be optimized.

The volume of EGR introduced into the combustion chamber is alsocontrolled to optimize the thermal mass of the combustion chamber basedon the signals identified above. The fine control provided by theengines of the present invention is not possible with external EGR heatexchangers, since the heat exchangers introduce a response lag into thesystem. In other words, adjustments made to the cooling of therecirculated exhaust gases at an external heat exchanger would not berealized in the combustion chamber until the exhaust gases in the heatexchanger finally arrive in the combustion chamber, which could takeseconds.

In an embodiment of the present invention, the inventive engine utilizesinternal EGR with direct cooling, as this provides the most immediateand precise control of EGR volume and exhaust gas temperature control.

Water injection volume and EGR volume is controlled based on pre-storedor periodically generated tables accessible by the engine controller. Inone embodiment, the tables are generated experimentally by runninginjection sweeps. Specifically, the engine is held at a constant speedand load while varying the amount of water injection and EGR. Theinjection sweeps are performed at various speeds and loads so that anoptimal value, or set of optimal values are identified for waterinjection and EGR under most operating conditions. Data is interpolatedbetween test results to produce a full matrix for the points that liebetween actual test points. Thus, the ECU is able to provide anoptimized water injection and EGR volume to the combustion chamber inorder to maintain desired operating parameters when the engine runsthrough various loads and speeds.

More specifically, a method 1000 for controlling the water and EGR foreach cylinder of an internal combustion chamber is described in FIG. 10.At 1010, the ECU determining current engine operating conditionsincluding, e.g., engine RPM, load, mass air flow. At 1015, the desiredair/fuel mixture is determined based on operating parameter such as themass air flow and RPM, for example.

The amount of EGR is obtained at 1020 based on the operating parametersas well as the air/fuel mixture. The amount of EGR may be obtainedempirically or based on a stored lookup table by the ECU. Additionally,the temperature of the exhaust gases is sensed in 1025 and reported tothe ECU.

Based on the air/fuel mixture, compression ratio and exhausttemperature, the necessary amount of cooling is calculated and theappropriate amount of water injection is determined in 1030 by the ECU.The amount of water to be injected may be empirically calculated ordetermined based on a pre-stored lookup table accessible by the ECU.

Based on the above determined values for air/fuel mixture, EGR level andWater injection volume, the ECU controls the fuel injector of thecurrent cylinder to inject air and fuel, at the calculated air/fuelratio, into the combustion chamber prior to top-dead-center (TDC) of thepiston in 1035. Additionally, at 1040 the water injector, andsimultaneously, at 1045 the EGR valve, are controlled to introduce thedetermined amounts of atomized water and exhaust gases into thecombustion chamber prior to TDC. In the present invention, the EGR valvemay constitute a valve disposed on an external EGR track, an exhaustvalve which is held open for a duration to allow exhaust gases torecirculate back into the combustion chamber, or an air intake valuecoupled to an EGR track, as described in greater detail above.

The atomized water and exhaust gases should be introduced at the sametime in order to induce more thorough mixing and cooling by the injectedwater, thus reducing the risk of premature ignition of the fuel in thecombustion chamber. Alternatively, the water and exhaust gases may beintroduced prior to introduction of the air/fuel mixture.

The ECU may continually monitor the performance of the engine and adjustthe values of water and EGR in their respective lookup tables.

That is, in one embodiment, using the predetermined information storedin one or more water injection and EGR tables, the engine controller,will compute the control parameters to affect the engine outputconditions such as the amount of atomized water and exhaust gases to beinjected into the combustion chamber. These adjustments are affected bythe engine controller communicating messages for controlling actuation(e.g., dwell time) of the fuel injector, communicating messages tocontrol the timing of water injection and the volume (before TDC) ofatomized water injection, and controlling the volume (before TDC) ofexhaust gases introduced into the combustion chamber, according to theembodiment described herein.

At an engine cycle-by-cycle basis, given the current sensed conditionsvalues, and in response to the current temperature and pressurereadings, and other variables, e.g., environmental conditions such asambient temperature, the engine controller will coordinate the operationof the system by sending out control messages for modifying the fuelinjection amount and timing, and control messages that control theamount of water injection (whether port or cylinder direct-injected)relative to the timing of the spark ignition (advance) at the cylinderduring the compression stroke for maximum efficiency, compression andcooling as described herein.

It is understood, that the monitoring and control of the engineoperations at any particular cycle of operation of the engine may beadjusted based on the operation during the prior cycle (including timeaverage of a few prior cycles) to ensure ignition and water injectionsoccurs at the proper crankshaft angle(s) in a stable manner.

Maintaining Engine Efficiency and Reducing NOx

In addition to the use of atomized water in the embodiments describedabove, an embodiment of the present invention is configured to inject aquantity of water into the combustion chamber of an internal combustionengine to maintain an engine temperature of between about 95° C. andabout 200° C. This temperature represents the exit temperature of thecoolant, i.e. radiator fluid, exiting the engine.

Introducing water into the combustion chamber prior to combustion of thefuel/air mixture can greatly reduce NOx. However, in a conventionalinternal combustion engine operating at a coolant temperature of about90° C., as the amount of water introduced increases, the efficiency ofthe internal combustion engine is decreased. The present inventionmaintains the efficiency of the internal combustion engine while greatlyreducing the generation of NOx emissions by operating the internalcombustion engine at a coolant temperature in the range of about 95° C.to about 200° C., and in another embodiment from about 100° C. to about200° C. and in still another embodiment, from about 100° C. to about140° C. Thus, in accordance with the present invention, the coolanttemperature can be 91° C., 92° C., 93° C., 94° C., 95° C., 96° C., 97°C., 98° C., 99° C., 100° C., 101° C., 102° C., 103° C., 104° C., 105°C., 106° C., 107° C., 108° C., 109° C., 110° C., 111° C., 112° C., 113°C., 114° C., 115° C., 116° C., 117° C., 118° C., 119° C., 120° C., 121°C., 122° C., 123° C., 124° C., 125° C., 126° C., 127° C., 128° C., 129°C., 130° C., 131° C., 132° C., 133° C., 134° C., 135° C., 136° C., 137°C., 138° C., 139° C., 140° C., 141° C., 142° C., 143° C., 144° C., 145°C., 146° C., 147° C., 148° C., 149° C., 150° C., 151° C., 152° C., 153°C., 154° C., 155° C., 156° C., 157° C., 158° C., 159° C., 160° C., 161°C., 162° C., 163° C., 164° C., 165° C., 166° C., 167° C., 168° C., 169°C., 170° C., 171° C., 172° C., 173° C., 174° C., 175° C., 176° C., 177°C., 178° C., 179° C., 180° C., 181° C., 182° C., 183° C., 184° C., 185°C., 186° C., 187° C., 188° C., 189° C., 190° C., 191° C., 192° C., 193°C., 194° C., 195° C., 196° C., 197° C., 198° C., 199° C., 200° C.

The breakdown temperature of the lubricants and seals used in the enginelimits the high end of the engine temperature. For example, conventionallubricants allow a high end temperature of about 140° C., whilesynthetic lubricants allow an upper range of at least 200° C. Thus, thepresent invention may be implemented using engine temperatures greaterthan 200° C. with the incorporation of lubricants, seals and otherengine components capable of properly operating at temperatures greaterthan 200° C.

For example, in an embodiment the engine shown in FIG. 5 is cooled byinternal water injection. The present embodiment includes all theelements shown in FIG. 5. However, in order to provide cooling for theinternal combustion engine, the quantity of water injected directly intothe combustion chamber 1 is controlled by the engine control computer 30based on the engine temperature.

The quantity of water injected into the combustion chamber 1 rangesbetween about 5% water to about 100% water with respect to the quantityof fuel being injected. In an embodiment, the quantity of water mayrange from about 25% to about 100% with respect to the quantity of fuel,provided the amount of injected water is higher than the volume of waterrequired to saturate air at room temperature.

The actual amount of water injected also depends on the location alongthe intake track 6. Thus, when the water injection occurs directly intothe combustion chamber 1, as shown in FIGS. 5, 6, 8 and 9, momentsbefore the piston 2 reaches top-dead-center, the quantity of waterinjected may be set at a lower end of the range, i.e., about 5% withrespect to the quantity of fuel injected, since the amount of elapsedtime between the water injection and combustion is shortened, resultingin less evaporation of the injected water. The longer the length of timethat elapses between the moment the water is injected into thecombustion chamber 1 and combustion of the fuel at top-dead-center thehigher the volume of injected water that will be needed.

The present embodiment is described above with respect to internalcombustion engines equipped with EGR. However the present embodiment maybe implemented in an internal combustion engine without EGR as well, asshown in FIG. 11. As shown in FIG. 11 the implementation is similar tothe embodiment shown in FIG. 5. However, the EGR track 11 and supportingcomponents are eliminated in this embodiment. The embodiments shown inFIGS. 6-9 may be similarly modified to forego the EGR component of thepresent invention and implement the internal cooling aspect of thepresent invention instead.

The system implementing the present embodiment, and shown in FIG. 12,includes an internal combustion engine 1102, a water reservoir 1104, awater injector 1106 arranged to inject water into an intake track orcombustion chamber of the internal combustion engine 1102, a fluid line1108 coupling the water reservoir 1104 and the water injector 1106, anda controller 1110 such as a microcontroller, CPU or FPGA configured toreceive operational data, such as exhaust temperature, reservoir watertemperature and engine rpm, from sensor probes 1112, and control theamount of water injected into the intake track or combustion chamber ofthe internal combustion engine 1102. The water reservoir 1104, waterinjector 1106 and fluid line 1108 form a water injection system.

The water reservoir 1104 is dimensioned to hold a sufficient volume ofwater for cooling the internal combustion engine 1102 for a determinedduration of time in operation or distance traveled. For example, if thequantity of water injected is set to equal the amount of fuel injected,then a vehicle with a 16 gallon fuel tank may be equipped with anequally sized water reservoir 1104, as well; thus allowing sufficientwater volume to cool the internal combustion engine 1102 for the fullrange of travel of the vehicle. Alternatively, a smaller water reservoir1104 may be provided, which would provide a more limited range oftravel, but reduce weight of the vehicle.

In an embodiment, the present invention may be configured to recoverwater vapor exiting through the exhaust track 1115 by way of a condenser1120 and water return line 1122 coupled between the exhaust track 115and the water reservoir 1104. In this embodiment, the range of travelcan be extended while still utilizing a small reservoir of only severalgallons.

In an embodiment, shown in FIG. 12, the present invention may beconfigured as the primary cooling system for the internal combustionengine 1102 comprising a water reservoir 1104, water injector 1106 and acondenser 1120. However, a secondary cooling system 1130 implemented asa conventional radiator and coolant reservoir system may be provided aswell.

It is understood that the figures show the implementation of embodimentsof the present invention with respect to an individual combustionchamber of an internal combustion engine for simplicity. However, inpractice the embodiments shown are implemented for each combustionchamber of an internal combustion engine. Thus, while one water injectoris shown and described throughout the figures and detailed description,it is understood that in embodiments where the water injector injectswater directly into the combustion chamber, at least a number of waterinjectors equal to the number of combustion chambers is provided.

In embodiments relating to water injection into the intake track, thepresent invention may be implemented with a single water injectordisposed before an intake manifold, which splits the intake intoindividual intake paths directed to each combustion chamber.Alternatively, if the water injector disposed after the intake manifold,at least one water injector is provided for each combustion chamber anddisposed in respective individual intake paths.

The secondary cooling system 1130 provides engine cooling by a flow ofcoolant, such as a glycol/water mixture, from the radiator throughcooling through-passages formed in the internal combustion engine 1102and returned to the radiator. At the radiator, the coolant is cooled inthe conventional manner using airflow generated either by movement ofthe vehicle or a fan. Thus, the two cooling systems may be configured tofunction in parallel. Alternatively the secondary cooling system 1130may be configured to function only once the water reservoir in theprimary cooling system has been emptied.

The benefits of the present invention are seen in a large reduction inaerodynamic drag resulting from the elimination of airflow through aradiator. Additionally, as shown in Table 2 below, an internalcombustion engine with a slightly elevated engine temperature of about130° C. allows operation using internal cooling with no net loss ofengine efficiency, while also reducing NOx emissions.

Experimental Data

The effect of internal cooling using water injection and bypassing theexternal heat exchanger circuit was studied for elevated coolanttemperatures. The test engine, operating at a constant 1800 rpm, wasmodified with a bypass circuit which directs the flow of coolant awayfrom the heat exchanger, i.e. radiator. This is done using a three waybypass valve. The coolant used for this test was formulated withoutwater to avoid boiling at the higher engine coolant temperaturesexperienced during the testing. Table 2 shows a summary of the results.

It can be seen that when the coolant temperature is maintained at 90°C., which is the coolant temperature set point for conventional internalcombustion engines, water injection provided by the present inventiongreatly reduces NOx, but at a cost of decreased engine efficiency. Forexample, at 6 bar, NOx was reduced by 91% with the internal coolingtemperature stabilized at 90° C. However, the engine efficiency was alsoreduced by 19%. By modestly increasing the coolant temperature to 130°C., engine efficiency is returned to 36%, while NOx generationexperiences only a slight 0.24 g/KWh increase to 1.22 g/KWh. Thus, itcan be seen that the increased coolant temperatures offset the adverseeffects caused by excessive water injection.

TABLE 2 Comparison of Conventional Cooling and the Present InventionCoolant Load Cooling Temp. BSFC Efficiency NOx EGT bar Method ° C. g/KWh% g/KWh ° C. 3 Conventional 90 269 31 10.7 291 3 Present Invention 90288 23 1.13 243 6 Conventional 90 218 36 11.7 378 6 Present Invention 90246 29 0.98 316 6 Conventional 130 225 37 12.5 441 6 Present Invention130 230 36 1.22 385

The described embodiments of the present invention are intended to beillustrative rather than restrictive, and are not intended to representevery embodiment of the present invention. Various modifications andvariations can be made without departing from the spirit or scope of theinvention as set forth in the following claims both literally and inequivalents recognized in law.

What is claimed is:
 1. An internal combustion engine comprising: atleast one cylinder, each cylinder having a combustion chamber, a piston,an air intake valve, and an exhaust valve, a mechanical compressionratio in each cylinder being greater than 12:1 and less than 40:1; anair intake track in communication with each air intake valve; an exhausttrack in communication with each exhaust valve; a fuel handling systemwith at least one fuel injector for injecting fuel into the combustionchamber or intake track, the fuel handing system providing an air tofuel ratio having a ratio of air to fuel (λ) greater than 1 and lessthan 7.0; an ignition system for igniting the fuel in the combustionchamber at an end portion of a compression stroke of the piston; a waterinjection system comprising a water injector for introducing an amountof water into the combustion chamber, and a water reservoir in fluidcommunication with the water injector, the water injector being arrangedto inject a controlled amount of liquid water stored in the waterreservoir into the combustion chamber or intake track; and an externalcooling system comprising a radiator and coolant, the external coolingsystem being configured to maintain a coolant temperature of betweenabout 91° C. and about 200° C.
 2. The internal combustion engine ofclaim 1, further comprising: an exhaust gas recirculating (EGR) systemfor recirculating exhaust gases from the exhaust port to the engineintake; and an EGR cooling system for cooling the recirculated exhaustgases by direct contact with a predetermined quantity of atomized waterinjected into the EGR track
 3. The internal combustion engine of claim1, wherein the amount of water is in a range of between about 5% andabout 100% with respect to the quantity of fuel injected.
 4. Theinternal combustion engine of claim 1, wherein the amount of water is ina range of between about 25% and about 100% with respect to the quantityof fuel injected.
 5. The internal combustion engine of claim 1, whereinthe internal combustion engine is maintained at a temperature of about130° C. during operation.
 6. An internal combustion engine comprising:at least one cylinder, each cylinder having a combustion chamber, apiston, an air intake valve, and an exhaust valve; an air intake trackin communication with each air intake valve; an exhaust track incommunication with each exhaust valve; a fuel handling system with atleast one fuel injector for injecting fuel into the combustion chamberor intake track; an ignition system for igniting the fuel in thecombustion chamber at an end portion of a compression stroke of thepiston; and a water injection system comprising a water injector forintroducing an amount of water into the combustion chamber, and a waterreservoir in fluid communication with the water injector, the waterinjector being arranged to inject a controlled amount of liquid waterstored in the water reservoir into the combustion chamber or intaketrack; and an external cooling system comprising a radiator and coolant,the external cooling system being configured to maintain a coolanttemperature of between about 91° C. and about 200° C.
 7. The internalcombustion engine of claim 6, wherein the amount of water is in a rangeof between about 5% and about 100% with respect to the quantity of fuelinjected.
 8. The internal combustion engine of claim 6, wherein theamount of water is in a range of between about 25% and about 100% withrespect to the quantity of fuel injected.
 9. The internal combustionengine of claim 6, wherein the internal combustion engine is maintainedat a temperature of about 130° C. during operation.